Machine tool transmission and controls therefor



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MACHINE TOOL TRANSMISSION AND CONTROLS THEREFOR Original Filed Aug. 1, 1955 14 Sheets-Sheet 1 INVENTOR.

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MACHINE TOOL TRANSMISSION AND CONTROLS THEREFOR United States Patent 3,027,782 MACHINE TOOL TRANSMISSION AND CONTROLS THEREFOR John C. Hollis, Fond du Lac, Wis., assignor to Giddlngs 62 Lewis Machine Tool Company, Fond du Lac, Wrs., a corporation of Wisconsin Continuation of appiication Ser. No. 525,469, Aug. 1, 1955. This application Aug. 31, 1960, Ser. No. 54,811 14 Claims. (Cl. 74-751) This invention relates in general to machine tools and in particular, to multi-speed transmission and their controls for driving massive rotatable elements of machine tools.

The general aim of the invention is to provide an improved multi-speed transmission which is especially adapted to drive massive machine tool elements while permitting changes in speed to be effected by controls which require only finger tip manipulation on the part of an operator.

An important object of the invention is to create such a multi-speed transmission which may be shifted to any one of a great number of drive ratios while the parts are in motion and under load.

In this respect, it is a related object to provide such an improved multi-speed transmission which eliminates severe impact upon the driven element as a result of speed changes, yet which achieves smooth speed transition without the necessity for a fluid or flexible coupling and the disadvantageous impositive drive inherent in such coupling.

Still another object of the invention is to provide an improved arrangement for braking the massive element driven by such a multi-speed transmission. More specifically, it is intended to create an arrangement in which braking elements are located remotely from the transmission components and the exterior of the transmission housing for oil-free operation and air cooling of braking surfaces, yielding ease in adjustment or replacement of the braking elements. The arrangement further contemplates brake means which work independently of the speed change gearing to provide uniform stopping action regardless of the speed setting and even in the event that one of the speed change clutches should fail.

A further object is to create finger tip controls for a multi-speed transmission through the provision of electric control circuitry and an improved, extremely compact rotary permutation switch.

A related object is to provide transmission controls for jogging the driven element together with means for assuring that the driven element is moved at its lowest speed when jogged, regardless of the prior speed setting of the transmission.

Another object is to provide a multi-speed transmission and controls for it which yield substantially all of the l advantages of constant cutting speed apparatus without requiring complex components and controls previously required to obtain those advantages.

Still another object is to provide a control arrangement in which the speed of a rotatable work support is varied in steps according to the positional zone of a tool holder teedable radially of the support.

Yet another object is to provide control means which automatically take into account the radius of a tool holder with respect to the axis of a rotatable work support, and which substantially maintain automatically any one of a plurality of cutting speeds selected in response to the setting of a dial by the operator.

Still further, it is an object of the invention to make the speed of a rotatable machine tool element controllable optionally by a manual selector, or by an auto- ICC matically actuated selector actuated according to the radial zone of a tool holder to provide substantially constant cutting speed, yet in which the manual selector takes control automatically if the operator moves it.

Other objects and advantages will become apparent as the following description proceeds, taken in conjunction with the accompanying drawings, in which:

FIG. 1 is a front elevation of an exemplary machine tool having a speed change transmission and controls embodying the features of the present invention;

FIG. 2 is a rear perspective of the machine tool shown in FIG. 1;

FIGS. 3 and 4, when joined along the lines X-X, are a longitudinal, vertical cross section of the multispeed transmission employed in the machine tool;

FIGS. 5 and 6 are sectional views taken respectively along the line 5-5 and 6-6 in FIG. 3;

FIG. 7 is a diagrammatic representation of the multispeed transmission illustrated by FIGS. 3-6;

FIG. 8 is a detail view illustrating the face of a control pendant on the machine tool shown by FIG. 1;

FIG. 9 is a detail view of a rotary selector switch shown in FIG. 8 and taken in section along the line 9-9 in FIG. 10;

FIGS. 10 and 11 are sectional views taken along the lines 10-10 and 11-11, respectively, in FIG. 9;

FIGS. 12-15 are detail views, in perspective, illustrating cams and switch trip plates associated therewith;

FIG. 16 is an enlarged front elevation of the right feed control on the machine tool illustrated in FIG. 1;

FIG. 17 is a sectional view taken substantially along the line 1717 in FIG. 16 and illustrating particularly means for actuating cutting speed controls;

FIG. 18 is a detail view of a trip ring shown in FIG. 17 for correlating the actuation of transmission controls with the radial zone of a tool holder;

FIG. 19 is a detail view taken substantially along the line 1919 in FIG. 17 and illustrating an adjustment scale permitting compensation for relative positions of a tool within the tool holder;

FIG. 20 is a vertical section, taken substantially along the line 20-20 in FIG. 21, illustrating the construction of a cutting speed selector and means for actuating it according to the position of a tool holder;

FIGS. 21 and 22 are detail sectional views taken substantially along the lines 21-21 and 22-22, respectively, in FIG. 20;

FIG. 23 is a diagrammatic representation of the cutting speed control illustrated in FIG. 20;

FIGS. 24 and 25 are graphs respectively, depicting the relative time efiiciency for machine operations obtained with theoretically perfect constant cutting speed control and with the constant cutting speed control of the present invention;

FIGS. 26 and 27 are electrical diagrammatic representations of the manual speed selector switch illustrated by FIGS. 9-15 and the cutting speed control switch illustrated by FIGS. 20-23, respectively;

FIG. 28 is a schematic diagram of the electrical circuit employed in the transmission controls; and

FIG. 29 is a diagrammatic illustration of the relationship of the multi-speed table drive, multi-speed saddle feed drive, and the saddle position indicator.

While the invention has been shown and is described in some detail with reference to a particular embodiment thereof, there is no intention that it thus be limited to such detail. On the contrary, the intention here is to cover all alterations, modifications, and equivalents falling within the spirit and scope of the invention as defined by the appended claims.

GENERAL DESCRIPTION OF THE EXEMPLARY MACHINE TOOL To make clear the environment of the preferred embodiment of the invention, a specific machine tool to which it is applied has been shown in FIGS. 1 and 2. As there illustrated, a large vertical turret lathe is provided With the improvements of the invention. The lathe itself includes a base or housing 31 extending rearwardly from a massive rotatable element or work support, in this instance, a work table 32 rotatable about a vertical axis and equipped with means such as chuck jaws 34 for holding any one of a variety of workpieces (not shown). Disposed within the housing 31 and extending in a generally fore-and-aft direction is the improved multi-speed transmission 35 (detailed in FIGS. 3-7) which drivingly connects a prime mover or main motor 36 with the table 32. The motor 36 may be of the constant speed induction type and the transmission 35 is effective to rotate the table 32 at any one of a plurality of speeds in response to the manual setting of a rotary selector 38 carried on a pendant 39.

The machine further includes a horizontal cross rail 40 which may be translated vertically along ways 41. A

tool holder, here a turret 42, is fixed to a ram 43 which may be fed vertically within a saddle 44, the latter being translatable horizontal along ways 45 of the cross rail 40, and thus substantially radially of the table 32. Vertical feed of the ram 43 and its position are indicated by a scale 46 on a right head 48, whilethe horizontal feed or position of the saddle 44 is indicated by a scale 49. The saddle feeding movement results from take-off of power from the multi-speed transmission 35 to a feed transmission 50a and thence to a lead screw 258a (FIG. 29). Feed motion at anyone of a plurality of rates (inches per table revolution) is selected by setting a rotary feed selector 50, as explained more fully in applicants copending application Serial No. 526,272, filed August 3, 1955, and issued as Patent 2,831,361 on April 22, 1958.

A hand wheel 51 permits manual feeding movements of the ram 43 and saddle 44.

In the present instance, the vertical turret lathe 30 is equipped with a left head similar to the right head 48 and controlling the vertical feed of a second tool holding ram 62 relative to a saddle 64 which is itself horizontally translatable along the crossrail 40. Still further, the'exemplary machine has a third alternative cutting tool holder 65 carried by a horizontally feedable ram 66 controlled by a head 6% which is translatable vertically along the ways 41 of the base 31.

A two-position manual lever 70 located at the right of the table permits an operator to select, as explaned below, table operation in either a high or low speed range.

THE MULTI-SPEED TRANSMISSION As shown best in FIG. 2, the main driving motor 36 is connected with the input end of the multi-speed transmission 35 by a plurality of V belts running over a relatively large sheave 81 which is disposed outside the rear wall of the main housing or casting 31. A removable belt guard may be employed to cover the V belts 81), but it permits free circulation of air around the sheave. Immediately above the sheave, as shown in FIG. 2 is located a control box 82 which may house the various relays and control components to be described later. With reference to FIG. 7, the multi-speed transmission 35 terminates in an output member or shaft 84 which carries a bevel gear 85 meshing with a ring gear 86 on the underside of the table 32 so that the latter is rotated about a vertical axis.

Connected between the sheave 81 as an input member and the shaft 84 asan output member are a plurality of planetary speed change gear sets 90, 91, 92 and 93 (FIGS. 3, 4 and 7) together with a fifth planetary gear set 94 which is controlled by the manual lever 70 (FIG. 1) to give high and low speed ranges. Each of these planetary gear sets -94 operates at either of two speed change ratios and, together, they make up the entire multi-speed transmission 35. With the manually controlled gear set 94 in the low range, the other four gear sets 90-93 thus make possible sixteen diflerent speed change ratios, assuming that the main motor 36 operates at constant speed. Shifting the gear set 94 to the high position permits sixteen additional table speeds, for a total of 32.

Referring more particularly to FIGS. 3-7 the first planetary speed change gear set 90 is made up of a first terminal or input member formed as an integral sun gear 1110 on the inner end of a sleeve shaft 101 which totatably surrounds a brake shaft 1512 (to be described) and which is keyed as at 104 to the sheave 81. The terminal output member of the gear set 90 is a planet carrier 105 which journals on studs 106a, a first plurality of planet gears 106 meshing with the sun gear 101) and, in turn, meshing with a second set of overlapping planet gears 108 journaled on studs 108a fixed to the carrier 1115. An intermediate or reaction member of the first gear set 90 is constituted by a gear 109 meshing with the planet gears 1113 at their inner sides. The reaction gear 109 .is integral With a sleeve 116 rotatable on the sleeve 101 and keyed as at 111 to a collar 112 having radially extending flanges 114, 115. If the collar 112 is held stationary by locking it to the housing 31, rotation of the input sun gear 101) makes the planet gears 106, 108 roll around the reaction gear 109, so that the output terminal member or planet carrier .105 rotates in the same direction as the input sun gear but at a reduced speed ratio. By way of example, the diameters of the several gears employed .may be selected to result in a speed reduction ratio of 2:1. On the other hand, if the collar 112 and the reaction gear 109 are locked fast to the input memher or sun gear 160 (i.e., locked to the sheave 81), the planet gears 106, 108 cannot rotate about their own axes and direct drive with a 1:1 ratio between the sheave 81 and the planet carrier 105 is obtained.

The second planetary gear set 91 also comprises an input terminal member, an output terminal member, and an intermediate or reaction member, shown, respectively, as a planet carrier 126, a ring gear 121, and a sun gear 122. The input terminal member or carrier is keyed fast at 124m a sleeve extension of the planet carrier 105 so that the gear sets 91 and 91 are connected in tandem relation. The carrier 120 supports a plurality of studs 125 journaling a like plurality of planet gears 126 which mesh at their outer sides with the ringgear 121 and at their inner sides with the reaction sun gear 122. From FIGS. 3 and 7 it will be seen that the reaction sun gear 122 is provided with a radial flange 127, and that the output ring gear 121 carries bolted thereto a radial flange 128 as well as a rotatable sleeve 129 formed with an integral sun gear 130 constituting the input terminal member for the next tandemly connected planetary gear set 92.

When the reaction sun gear 122 and its radial flange 127 are held stationary with the housing 31, and the planet carrier 120 is rotationally driven, the planet gears 126 must rotate about their studs 125 and roll around the sun gear so as to drive the ring gear 121 at an increased speed. The diameters of the several gears may, for example, be chosen so that the output ring gear 121 turns 1.4 times faster than the input planet carrier 1212, and in the same direction. On the other hand, if the reaction sun gear 122 and its radial flange 127 are held fast to the input terminal member or carrier 120 (i.e., to the planet carrier 105), then the planet gears 126 cannot rotate and direct drive with a 1:1 ratio is obtained between the input terminal member 129 and the output terminal member 121.

Deferring a detailed description of the third planetary gear set 92 for the moment, it may be observed that this gear set has in addition to the input terminal member formed by the sun gear 130, an output terminal member formed by a second sun gear 14 1 The latter is made integral with or keyed to a hollow sleeve 14]. (FIG. 4) which at its opposite end carries an input terminal member or sun gear 142 for the fourth planetary gear set 93. This fourth gear set includes as its output terminal member a planet carrier 144 having keyed to it a hollow sleeve 145 extending axially to the left and, in turn, keyed to a sun gear 146 forming the input terminal member of the next tandernly connected planetary gear set Finally, the fourth gear set 93 employs a reaction memher or ring gear 143 which meshes with the peripheries of a plurality of planet gears 149 journaled on studs 15% supported by the planet carrier 14 4. The planet gears 149 also mesh with the input sun gear 142.

If the reaction ring gear is held stationary with the housing 31, rotation of the input sun gear 142 forces the planet gears 149 to rotate about their studs 1559, thereby driving the output planet carrier ass in the s me direction but with a reduced speed ratio. By way of example, the diameters of the several gears may be chosen to make this speed reduction ratio in the order of 4:1. However, if the planet carrier 144 is held rigid with the input sun gear 142, then the planet gears 149 cannot rotate about their own axes and a direct drive with 1:1 ratio results.

As noted above, the sun gear 146 which is rotationally rigid with the planet carrier 144 forms the input termi nal member for the fifth speed change gear set 94. This gear set also includes an output terminal member constituted by a planet carrier 155 which is integral with the output shaft 84. An intermediate member or internal ring gear 156 meshes at the outer peripheries of a plurality of planet gears 158 journaled on studs 159 supported by the planet carrier 155. These planet gears also mesh at their inner sides with the sun gear 146.

With this arrangement, therefore, if the reaction ring gear 156 is held stationary with the housing 31, the planet gears 158 must roll around the ring gear 156 and gear 146 so that the planet carrier 155 turns at a reduced speed relative to that of the input sun gear M6. in an exemplary installation, the relative diameters of the several gears may be chosen such that this speed reduction ratio is in the order of 4:1. On the other hand, if the ring gear 156 is held rigid with the input sun gear 146, i.e., locked to the sleeve 145, then the planet gears 158 cannot rotate about their axes and a direct lzl drive ratio between the sun gear 146 and the planet carrier 145 results.

Impact Minimization From the foregoing, it will be apparent that by locking the reaction member of each gear set either to the housing or to one of the terminal members in different combinations, a total of sixteen speeds may be obtained for each setting of the fifth gear set 9 4. The gear set 92, to be described in detail below, produces a speed ratio change in a manner similar to that of the other gear sets, but it is constructed in a particular manner to minimize the impact or jolt which would otherwise be given to the massive work table by a transition in one or more of the gear sets while operating under load.

In planetary transmissions of the type here considered it has in the past been deemed necessary to include in the drive path a flexible or fluid coupling which absorbs the impact created by clutching or braking of rotating parts while the driven element was in motion. Such a fluid coupling renders the drive impositive and totally nosatisfactory for application to machine tools where the speed of a rotatable work support must be accurately controlled for thread cutting and certain other operations. Prior arrangements without a flexible coupling resulted not only in shocks on the driven element, but also placed tremendous loads on the prime mover when the various clutches were shifted to create an increase in table speed. This dictated the use of a motor having a much higher horse power rating than that which would be required under steady state operation.

In accordance with an important feature of the invention, these difficulties are eliminated and a smoothly shiftable tandem planetary speed change transmission achieved, without the use of a fluid coupling, by constructing one of the planetary gear sets in a manner such that its reaction member tends to rotate at a high speed while being shifted from the housing to a terminal elelent, thereby storing kinetic energy which is, in part,

returned to the transmission as the speed transition is completed. Moreover, the controls are arranged such that this particular kinetic energy storing planetary gear et is shifted each time that a speed change is made.

Referring in more detail to FIGS. 3, 6 and 7, the planetary gear set 92 which performs this energy storing function has the sun gears and 14% as its input and out put terminal elements, respectively. This gear set further includes an intermediate member 178 formed as a relatively long internal ring gear which meshes around the outer eripheries of two sets of planetary gears 171 and 172, respectively. The latter gears are journaled on axial studs 173 carried by a freely rotatable planetary carrier 174. It will be noted from FIG. 3 that the reaction ring gear 174} has an integral radial flange 17% which, as explained below, may be held stationary with the housing or rotationally rigid with the input sun gear 13th or, for the same effect, the preceding output ring gear 121.

With the arrangement shown, the input sun gear 13% is larger in diameter than the output sun gear 149 and the planetary gears 171 are smaller in diameter than the planet gears 172. With the reaction ring gear 17% held fast to the housing, the gear set 92 behaves as a pair of gear sets connected in tandem, one providing a speed step-down and the other providing a speed step-up. The planet gears 171 in rotating about their respective axes as the input sun gear is driven, cause the common planet carrier 174 to rotate at a lower speed than that of the input sun gear. The carrier 174 in bodily rotating the planet gears 172 causes them to roll around the internal teeth of the ring gear 17% so that the output sun gear is driven at a greater speed than the planet carrier itself.

In the present instance, the relative diameters of the several gears are so chosen that the net speed change or step up, between the input sun gear 1% and the output sun gear 140, with the ring gear held stationary, is in the order of 1.2: 1.

On the other hand, if the ring gear 17th and its flange 170a is locked to the input sun gear 130, then the planetary gears 171 and 172 cannot rotate, and the two sun gears 130, 14! are driven in unison with a one-to-one ratio.

During the transition period when the ring gear 170 is being shifted from the housing to a part rigid with the input sun gear, that ring gear is free to rotate. And by virtue of the two planet gear groups 171, 172. and a floating planet carrier 174, the ring gear 17%) tends to rotate in the same direction as the input sun gear 130 if the latter speeds up faster than the sun gear let out in the opposite direction if the sun gear 130 lags the sun gear 140. The ring gear 17h thus rotates in opposite directions when the speed change is an increase or a decrease, respectively. Moreover, the relative diameters of the gears cause the ring gear 17% to attempt to rotate at a speed about six times the difference in the speeds of the two sun gears 13% and 14%). The kinetic energy put into the ring gear is thus fairly great, but proportional to the difference in speed between the motor 36 and the table 32.

As the ring gear 179 thus tends to rotate rapidly during the time that it is free, it stores a large amount of kinetic energy, especially in view of its relatively large diameter and its radially extending flange 170a. Thus, the prime mover or motor 36 is temporarily loaded and its energy stored in the rotating ring gear 171) even though it is momentarily disconnected from the table 32. The motor cannot accelerate rapidly only to be abruptly loaded again. A heavy impact, which would abruptly load the motor is prevented as the ring gear 17 is re-locked since its energy is in part returned to the high inertia work table, aiding in accelerating the same.

The result is a smooth impulse, as opposed to an abrupt impact, imparted to the relatively light ring gear 170 which can, owing to its smaller inertia, speed up quickly to store energy in kinetic form. Then, as the ring gear 170 is re-locked, by friction means described, its energy is in part dissipated as heat and the remainder is smoothly transferred through the succeeding gear sets 93, 94 to the table 32. The total effect of this arrangement is such that the massive table 32 may be given changes in speed without heavy impacts on the gearing and without employing any type of fluid coupling.

As an important aspect of the invention, the energy storing gear set 92 is shifted each time the table speed is changed, as explained more fully below. Whenever any of the other gear sets is shifted, the gear set 92 is also shifted so that the smoothing energy storage action occurs.

Electric Brakes and Clutches For selectively holding the reaction members of each gear set either stationary with the housing 31 or rotationally rigid with one of their corresponding terminal members, electro-magnetic brakes and clutches are preferably employed. As shown in FIGS. 3, 4 and 7, the reaction gear 109 of the gear set 91 may be held stationary with the housing 31 by energizing a brake coil 96b which magnetically attracts the flange 114, slidable on pins 114a carried by the collar 112, against a coil shoe 11 1b mounted rigidly in the housing 31. This as explained, causes the first gear set 90 to produce a speed reduction of 2:1 between the sheave 81 and the terminal member or planet carrier 105. Alternatively, a clutch coil 990 may be energized to magnetically attract an armature 81a, slidable on pins 81b carried by the sheave 81, against a friction surface on the flange 115 which, in turn, is bolted to the collar 112. This, of course, results in a direct or one-toone drive through the first gear set. The flange 115 has a recess into which the coil 913s extends, but the coil is stationary while the flange rotates.

In a similar manner, the reaction gear 122 may be held stationary by energizing a brake coil 9112 which attracts an armature 127a, carried by pins 12717 in the flange 127, against a friction surface 91a on a shoe mounted rigidly in the housing 31. This produces the 1.4-:1 speed step-up between the input planet carrier 120 and the output ring gear 121. Alternatively, the brake 9111 may be de-energized and a clutch coil 91c energized to attract a second armature 127a (also carried on the pins 12%) against a friction surface 91d on a ring 1115a which is bolted fast to the ring gear 105 and rotated in closely spaced relation to the coil 910. The energization of the clutch coil 91c thus locks the reaction gear 122 to the input member 120 and produces a direct drive through the second gear set 91.

An electric brake coil 92b, and an electric clutch coil 920 are likewise employed to hold the reaction ring gear 170 of the gear set 92 either stationary or rotationally rigid with the input sun gear 131 which, in turn, is rotationally rigid with the ring gear 121. As shown in FIGS. 3 and 4, energization of the brake coil 92b attracts an armature 17Gb carried on pins 1700 in the flange 170a against a friction surface 92a of a shoe which surrounds coil 92b and is mounted fast in the housing 31. Conversely, energization of the coil 92c attracts an armature 179d against a friction surface 128a on the flange 128 bolted fast to a radial projection from the axial sleeve 129 which is rotationally rigid with the ring gear 121.

For locking the reaction ring gear 14-8 of the gear set 93 stationary with the housing 31, a multiple disc electromagnetic clutch is employed having a brake coil 93b and a set of interleaved friction discs 93a, 930! which are carried respectively by an axial sleeve 148a fixed to the ring gear 148, and a surrounding collar bolted fast to a partition 1&1 in the housing 31. This, as explained, produces a speed reduction ratio of 4:1 between the input sun gear 142 and the output planet carrier 144. For locking the gear set 93 so that it produces direct drive, a multiple disc electro-magnetic clutch is employed which includes a clutch coil 93c and a plurality of interleaved clutch discs 93e, 93f. The discs 93a are rotationally fast, yet accurately shiftable, in a ring 144a rigid with the planet carrier 144, while the discs 93 are similarly mounted in a sleeve 141a fast on the sleeve 141, which in turn is connected to the input sun gear 142.

The fifth gear set 94 is shifted to produce either of its two speed change ratios by manipulation of the lever '70 (FIG. 1). An appropriate linkage is made between that lever and a shifter yoke (not shown) engaged in a circumferential slot 156a of the ring gear 156. Shifting of the ring gear 156 to the left (FIG. 4) causes its inner teeth tolock with mating clutch teeth' on the collar bolted fast in the housing 31, these causing the gear set 94 to provide, in the present instance, a 4:1 speed reduction between its input sun gear 146 and its output carrier 155. On the other hand, the ring gear 156 may be shifted axially to the right so that its internal teeth clutch with mating teeth 1556 on a clutch flange 1% locked rigidly on the shaft 145 which turns with the input sun gear 146. This produces direct drive through the gear set 94.

While it is possible to vary the specific speed change ratios provided by a particular transmission constructed in accordance with the invention, the following table gives exemplary table speeds obtained by energizing the several brake coils and clutch coils in various combinations, assuming that the main motor 36 operates at a constant speed of about 1725 revolutions per minute.

Brake and Clutch Pattern; Table Speed x=energized, o=deenergized r.p.n1.

Pattern Number High Low 92b 92c 91!? 91c 90c 90b 93:- 93b Range Range 1:1 in 4:1 in

set 94 set 94 o x 0 x o x o x 24 6.0 x o 0 x o x o x 28 7.0 o x x o o x 0 x 34 8. 4 x o x o o x 0 x 40 10 o x o x x o 0 x 48 12 x o 0 x x o 0 x 56 14 o x x o x o 0 x 68 17 x 0 x o x o 0 x 80 2O 0 x o x o x x o 96 24 x 0 o x o x x o 112 28 o x X 0 0 x x 0 136 34. x o x o o x x 0 160 40 o x o x x o x o 192 48 x o o x x o x o 224 56 o x x o x 0 x o 272 68 x 0 x o x 0 x o 320 80 T he Braking Arrangement As shown best by FIGS. 3 and 4, the table 32 is braked, when it is to be stopped, by an electro-rnagnetic brake which is located externally of the housing 31 and exposed at the rear of the latter. For this purpose, each of the axially alined, tandemly connected gear sets 90-94 is formed with a central axial passage therethrough, defined by the inner surfaces of the hollow shafts 101, 141, and 145. The brake shaft 102 extends through these passages and lies on the axis which is common to all of the gear sets. At its forward end, this brake shaft is keyed as at 1% to the sun gear 146 of the fifth planet gear set 94. At its rear end, the brake shaft 102 is journaled in bearings 191 and is keyed as at 1112 to a radial flange 194 which supports an armature 195 axially shiftable on pins 196. The main brake includes a stationary shoe 197 locked to a plate 198 which is supported by spacers 199 and bolts 2% from the rear surface of the housing 31. The stationary brake shoe 197 nestingly receives an electro-magnetic brake coil 201 and has a friction surface 197a against which the armature 195 is attracted when the coil 201 is energized.

During braking, at least one of the reaction members in the multi-speed transmission is left in neutral by having both its clutch coil and its brake coil de-energized. At the same time, the brake coil 201 is energized to frictionally lock the armature 195 and the flange 194 against rotation, so that the brake shaft 102 exerts a retarding torque on the sun gear 146 at the front of the transmission, thus braking the table 131 through only the fifth gear set 94.

Several important advantages accrue from this braking arrangement. First, the brake components such as the coil 231 the stationary shoe 197, and the rotatable armature 195 are all located externally of the housing 31 where they are free of lubricant which is supplied to the gears of the multi-speed transmission. Operation of frictional braking surfaces dry permits them to have a much greater torque capacity for a given size and area. Secondly, these brake components are readily cooled by air convection since any protective belt guard may be sufliciently open to permit good air circulation, in contrast to the relatively air-tight housing 31 which must be sealed against oil leakage. Moreover, these brake parts are readily accessible for adjustment, repair, or replacement. And when the brake acts, it completely by-passes the gearing in the first four gear sets 90-93, eliminating a wind-up and positively braking the table even if one of the several electro-magnetic clutches or brakes employed in those gear sets should fail. An additional advantage lies in the fact that the main brake always produces the same action in stopping the table. If the brake acted through the transmission components it would retard the table more abruptly when the transmission were set to produce a low speed and thus operated with a high mechanical advantage. If the brake were sufficiently large to stop the table in the desired minimum time when the transmission was set for a high table speed (low mechanical advantage), then it would be over-sized when the transmission was set for a low speed, stopping the table with a severe jolt. That difficulty is eliminated with the present compact arrangement.

TABLE SPEED SELECTOR As previously indicated, the main motor 36, in the interest of economy, may be of the induction type which runs at substantially constant speed. By energizing the various brake coils 90b-93b and clutch coils 900-930 in different combinations, sixteen speeds for the table 32 may be obtained in both the high and low ranges.

In accordance with one feature of the invention, the controls for obtaining such a plurality of speeds includes a novel permutation switch assembly for speed selection which is extremely small and compact and which may be easily set to any position by finger-tip adjustment. To achieve such compactness, a plurality of miniature, pressure-sensitive switches of the type commercially available under the trade names Micro-Switch and Switchette are mounted with their actuators in a common plane, angularly spaced about a rotatable shaft carrying a plurality of radial cams spaced axially thereon. The cams have a progressively increasing number of radial transitions increasing by powers of two, and each is engaged by a follower tab staggered axially on a different one of a plurality of pivoted trip plates, each of which has operative engagement with the actuator of a corresponding one of the switches.

Referring in more detail to FIGS. 8-l5, the manual speed selector 38 (mounted in the pendant 39) is made as a permutation switch assembly which comprises a frame made up of front and rear apertured plates 210, 211 which are held together by a plurality of fasteners 212. A shaft 214 is extended through the frame plates and rotatably journaled therein, having a square cross-sectional portion 10 214a disposed between the plates. A suitable face plate 215 may be fixed to the frame plate 210 and marked with numerical indicia indicating the table speeds as shown in FIG. 8 according to the rotational position of a mark on a hand knob 216 locked by a set screw 218 on the shaft and used to index the latter.

A plurality of switches 51, S2, S3, and S4 are disposed between plates 210, 211 in angularly spaced relation about the shaft 214. The fasteners 212 which hold the plates 211}, 211 together may be passed directly through mounting holes in the switch bodies, the latter thus serving to space the plates apart. The several switches, therefore, are disposed in substantially the same plane and present their respective yieldable actuators S111, 52a, 83a, and 84a angularly spaced about the shaft and extending radially inward toward the latter. Each of the switches contains normally open and normally closed contacts which are reversed when their respective actuators are depressed, as explained more fully below.

To produce every possible combination of switch actuation, a plurality of cams C1, C2, C3, and C4, are mounted in axially spaced relation on the square portion 214:: of the shaft 214. The cams have center apertures 22d matching the cross-section of the shaft portion 214a so that they must rotate with the latter. It will be observed from FIGS. 12-15 that the cam C1 is formed with tWo radial transition points 221, 222 between relatively large and small radius peripheral portions, these transition points being spaced 180 apart. The second cam C2 has four such radial transition points 2244 .27 between four peripheral portions two of which are of relatively large radius and two of which are relatively small in radius, all subtending approximately The third cam C3 has eight radial transition points 228 spaced at 45 angles and joining eight peripheral portions of alternately large and small radius. Finally, cam C4 has sixteen radial transition points 229 spaced apart 22 /2 and' separating sixteen peripheral portions of alternately large and small radius. The cams are phased" on the shaft portion 214a so that a large radius portion of each lies at the same angle.

For operatively associating each of the cams with a 7 corresponding one of the switches, a plurality of trip plates 230, 231 232 and 233 are pivotally mounted on a corresponding plurality of collared screws 234 which may also serve to mount a fifth switch I as explained below. Each of the trip plates extends from its pivot mounting to overlie a corresponding one of the switch actuators SlaS4a, the plate being substantially equal in axial length to the space taken by the several cams C1C4. To make each trip plate responsive only to a corresponding one of the cams, the plates are each provided with an inturned follower tab 23tla-233a, respectively. The follower tabs are progressively spaced axially along the shaft (compare FIGS. 12-15) so that the tab 2311a rides on the cam C1 and causes the trip plate 230 to actuate the switch S1, the tab 232a rides along the cam C2 and causes the trip plate 232 to actuate the switch S2, the tab 231a rides along the cam C3 and causes the trip plate 231 to actuate the switch S3, and the tab 233a rides on the cam C4 causing the trip plate 233 to actuate the switch S4.

The face plate 215 carries indicia at sixteen angularly spaced positions (FIG. 8). To hold the shaft 214- and the several cams at any one of these positions, after they are set by manually adjusting the knob 216, the cams C1-C4 are provided with small axial holes 235 which are alined after the cams are assembled on the shaft. A compression spring 236 disposed within the passageway formed by these alined holes (FIG.ll) urges a pair of detent balls 238 outwardly into yieldable engagement with corresponding ones of a plurality of sixteen angularly spaced detent holes 239 drilled in the front and rear frame plates 2.10, 211.

To afford utmost convenience to jogging controls, the shaft 214 is made hollow with a central passageway 1 1 214b therethrough. A plunger 242 is axially slidable in the passageway 2114b as a push button 244- in the center of the knob 215 is depressed. The plunger 242 is operatively connected to depress the actuator 245a of a fifth miniature switch I mounted by a suitable bracket to the rear of the frame plate 211. The switch J, when intermittently actuated causes the table 32 to jog, as described more fully below.

From the foregoing, it will be seen that the switch S4- is actuated or released each time that the knob 216 and the shaft 214 are indexed from one position to the next. That is, the switch S4 is actuated in the eight alternate positions of the sixteen possible positions for the knob 216. Since the cam C3 has eight radial transition points 228 and correspondingly only four peripheral portions of relatively great radius, the switch S3 is actuated or deactuated each time that the shaft 214 is indexed through two successive rotational positions. That is, the switch S3 is actuated when the shaft is in every other pm'r of its sixteen successive positions. Similarly, because the cam C2 has only two peripheral portions of relatively great radius, the switch S2 is actuated when the shaft is in alternate groups of four of the possible sixteen positions. Finally, the switch S1 is actuated when the shaft 214 is in eight of the sixteen possible positions.

It is clear, therefore, that the conditions of the switches will change in a permutated pattern as illustrated by the following table:

Switch Condition: x=Actuated, o= Deactuated Speed Indication,

4 NP KNP OOOOOOOO NNNNOOOONNHNOOOO WNOONNOOKMQONNOO NONONONOP'IQNQNONQ Referring to FIG. 26, the four switches 51-8 are diagrammatically represented to indicate that each has a pair of normally open contacts designated by the second numeral 1 and a pair of normally open contacts designated by the second subscript 2. Taking the switch S1. as an example, the contacts S14 will be closed, and the contacts S1-2 will be open when that particular switch is actuated.

It is to be understood from the foregoing that a permutation switch assembly such as that described may be constructed to employ any number of individual switches with a corresponding number of cams. if a plurality of switches S1, S2 Sn are employed, then a corresponding plurality of cams C1, C2, Cn will be used, each of the cams having a number of radial transitions and a number of alternately large and small radius peripheral portions equal to 2 raised to the power of the subscript Which designates that particular cam. That is, the number of transitions on the nth cam would equal 2. All of the switches may be mounted in angularly spaced relation about the cam-carrying shaft and provided with trip plates engaging the actuator of the corresponding switch, together with a follower tab riding on a corresponding one of the cams. For example, if a fifth switch and a fifth cam were added to the assembly shown in FIGS. 8l.5, then it would be possible to energize the five switches in a total of thirty-two possible combinations. In that case, of course, the detent means would be constructed to hold the shaft at any one of thirty-two angular positions and 12 the fifth cam would have thirty-two transitions between relatively short and long radial portions.

The connections of the several switch contacts described above into control circuit of the multi-speed transmission will be detailed below.

CUTTING SPEED CONTROL MEANS As indicated above, the saddle 44 is feedable horizontally and radially relative to the table 32 along ways 45 on the rail 40 (FIG. 1). The power for such feeding movement may be taken by a gearing connection from the multi-speed transmission itself. The rate of such feeding movement is controlled by the rotary feed selector 5i and the direction of such feeding movement by a swivel direction selector lever 250 as more fully described and claimed in applicants copending US. application Serial No. 526,272, filed August 3, 1955, and issued as Patent No. 2,831,361 on April 22, 1958 (as per admt. 6/ 19/59).

it has been recognized in the past that the maximum speed with which any machining operation may be successfully carried out is dependent upon the material and characteristics of the particular workpiece and the particular cutting tool being used. It is important both in the interest of efiiciency and uniform finish on the workpiece to maintain the cutting speed, i.e., the rate at which the cutting tool moves relative to the surface of the workpiece in a cutting direction substantially constant and near the maximum permissible value. This objective has been met in the past by relatively complex gearing or electronic controls which, in theory at least, maintain the cutting speed of a tool along a spiral path almost perfectly constant as it advanced radially across a workpiece. In other words, the speed of the work table was smoothly increased as a tool was fed inwardly from the periphery toward the center of the workpiece. Most common means for effecting this as smooth change in table speed included a variable speed motor drive together with elaborate motor speed control circuitry. Such equipment is, of course, quite expensive, and often fails to produce the theoretically perfect result of an absolutely constant cutting speed.

In accordance with an important aspect of this invention, means are provided, in combination with a multispeed transmission, for automatically varying the speed of a rotatable work support in steps according to zones of radial distance of the feedable tool or tool holder from the axis of the work support. The speed of the table increases as the tool holder is fed radially inward of the table so that a substantially constant cutting speed is maintained.

In carrying out the invention, a second permutation switch assembly 38A (FIG. 21), substantially identical to that described in connection with FIGS. 8-15 except for the omission of the manual knob 216, is employed. Such second switch assembly is connectable in control of the multi-speed transmission clutches and brakes in lieu of the first. It is positioned or stepped progressively in response to movement of the tool holder from one radial zone to the next. Additionally, a manual knob is connected with the second switch assembly to permit adjust ments in the particular value of the cutting speed which is substantially maintained by this step control.

In the preferred form, the actuation of the cutting speed control means is derived from the position of an indicator scale which moves in proportion to feed movements of a tool holder. As shown in FIG. 17, the indicator scale 49 for the saddle 44 is made as an angular ring and connected by gears 255, 256 with a shaft 258. The gears 255 and 256 form part of a high ratio reduction mechanism such that the scale 49 moves through one revolution as the shaft 258 makes many turns. The indicator scale arrangement includes a stationary index ring 259 and a micro-dial 26th, the whole being constructed and arranged as described in applicants copending applica- 

